ABAQUS Lecture Lecture Notes
By: Mohammad Javad Kazemzadeh-Parsi Assistant professor of Mechanical Engineering Islamic Azad University, Shiraz Branch Shiraz, Iran
ABAQUS Lecture Notes
M.J. Kazemzadeh-Parsi Kazemzadeh-P arsi
Chapter 1 Introduction The finite element method is a numerical method that can be used for fo r the accurate solution of complex engineering problems. Although the origins of the method can be traced to sev eral centuries back, most of the computational details have been developed in mid 1950s, primarily in the context of the analysis of aircraft s tructures. Thereafter, within a decade, the potential of the method for the solution of different types of applied science and engineering problems was recognized. Over the years, the finite element technique has been so well established that today, it is considered to be one of the best methods for solving a wide variety of practical problems efficiently. In addition, the method has become one of the a ctive research areas not only for engineers but also for applied mathematicians. One of the main reasons for the popularity of the method in different fields of engineering is that once a general computer program is written, it can be used for the solution of a variety of problems simply by changing the input data. The ABAQUS finite element software has strong s trong capabilities for solving, specifically, nonlinear problems and was developed by Hibbitt, Karlsson&Sorenson, Inc. The solution of a general problem by ABAQUS involves three stages: ABAQUS Preprocessor, ABAQUS Solver, and ABAQUS Postprocessor. ABAQUS/CAE or another suitable pre-processor provides a compatible input file to ABAQUS. ABAQUS/Standard or ABAQUS/Explicit can be used as ABAQUS/Solver to solve the problem. The ABAQUS/Standard, based on implicit algorithm, is good for static, strongly nonlinear problems. ABAQUS/Explicit, based on explicit algorithm, is intended for dynamic problems. Both ABAQUS/Standard and ABAQUS/Explicit can be executed under ABAQUS/CAE. The ABAQUS/CAE or another suitable postprocessor can be used for displaying the output (results) of the problem. ABAQUS/CAE provides a complete ABAQUS environment that provides a simple, consistent interface for creating, submitting, monitoring, and evaluating results from ABAQUS/Standard and ABAQUS/Explicit ABAQUS/Explicit simulations. ABAQUS/CAE is divided into modules, mod ules, where each module defines a logical aspect of the modeling process; for example, for defining the geometry, defining the material properties, and generating a mesh. As we move from one module to another module, we build the model from which ABAQUS/CAE generates an input file that we can submit to the ABAQUS/Standard or ABAQUS/Explicit for carrying the analysis. After completing the analysis, the unit (ABAQUS/Standard or ABAQUS/Explicit) sends the information to ABAQUS/CAE to allow us to monitor the progress of the job, and generates an output database. Finally, we use the visualization module of ABAQUS/CAE (also licensed separately as ABAQUS/Viewer) to read the output database and view the results of analysis. The ABAQUS/Viewer provides graphical displays of ABAQUS finite element models and results. It obtains the model and results information from the output database. We can control the output information displayed. For example, we can obtain plots such as undeformed shape, deformed shape, contours,x-ydata, and time history animation from ABAQUS/Viewer.
Basic Concepts of the Finite Element El ement Method The basic idea in the finite element method is to find the solution of a complicated problem by replacing it by a simpler one. Since the actual problem is replaced by a simpler one in finding the solution, we will be able to find only an approximate solution rather than the exact
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solution. The existing mathematical tools will not be sufficient to find the exact solution (and sometimes, even an approximate solution) of most of the practical problems. Thus, in the absence of any other convenient method to find even the approximate solution of a given problem, we have to prefer the finite element method. Moreover, in the finite element method, it will often be possible to improve or refine the approximate solution by spending more computational effort. In the finite element method, the solution region is considered as built up of many small, interconnected subregions called finite elements. As an example of how a finite element model might be used to represent a complex geometrical shape, consider the milling machine structure shown in Figure 1-1(a). Since it is very difficult to find the exact response (like stresses and displacements) of the machine under any specified cutting (loading) condition, this structure is approximated as composed of several pieces as shown in F igure 1-1(b) in the finite element method. In each piece or element, a convenient approximate solution is assumed and the conditions of overall equilibrium of the structure are derived. The satisfaction of these conditions will yield an approximate solution for the displacements and stresses. Figure 1-2 shows the finite element idealization of a fighter aircraft.
Figure 1-1
Figure 1-2
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Historical Background Although the name of the finite element method was given recently, the concept dates back for several centuries. For example, ancient mathematicians found the circumference of a circle by approximating it by the perimeter of a polygon as shown in figure 1-3. In terms of the present-day notation, each side of the polygon can be called a “finite element” . By considering the approximating polygon inscribed or circumscribed, one can obtain a lower bound S(l) or an upper bound S (u) for the true circumference S. Furthermore, as the number of sides of the polygon is increased, the approximate values converge to the true value. These characteristics, as will be seen later, will hold true in any general finite element application.
Figure 1-3
To find the differential equation of a surface of minimum area bounded by a specified closed curve, Schellback discretized the surface into several triangles and used a finite difference expression to find the total discretized area in 1851. In the current finite element method, a differential equation is solved by replacing it by a set of algebraic equations. In 1943, Courant presented a method of determining the torsional rigidity of a hollow shaft by dividing the cross section into several triangles and using a linear variation of the stress function ϕ over each triangle in terms of the values of ϕ at net points (called nodes in the present day finite element terminology). This work is considered by some to be the origin of the present-day finite element method. Since mid-1950s, engineers in aircraft industry have worked on developing approximate methods for the prediction of stresses induced in aircraft wings. In 1956, Turner, Cough, Martin, and Topp presented a method for modeling the wing skin using three-node triangles. At about the same time, Argyris and Kelsey presented several papers outlining matrix procedures, which contained some of the finite element ideas, for the solution of structural analysis problems. The name finite element was coined, for the first time, by Clough in 1960. Although the finite element method was originally developed mostly based on intuition and physical argument, the method was recognized as a form of the classical Rayleigh-Ritz method in the early 1960s. Once the mathematical basis of the method was recognized, the developments of new finite elements for different types of problems and the popularity of the method started to grow almost exponentially. The digital computer provided a rapid means of performing the many calculations involved in the finite element analysis and made the method practically viable. Along with the development of high-speed digital computers, the application of the finite
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element method also progressed at a very impressive rate. Zienkiewicz and Cheung presented the broad interpretation of the method and its applicability to any general field problem. With this broad interpretation of the finite element method, it has been f ound that the finite element equations can also be derived by using a weighted residual method such as Galerkin method or the least squares approach. This led to widespread interest among applied mathematicians in applying the finite element method for the solution of linear and nonlinear differential equations. It is to be noted that traditionally, mathematicians developed techniques such as matrix theory and solution methods for differential equations, and engineers used those methods to solve engineering analysis problems. Only in the case of finite element method, engineers developed and perfected the technique and applied mathematicians use the method for the solution of complex ordinary and partial differential equations. Today, it has become an industry standard to solve practical engineering problems using the finite element method. Millions of degrees of freedom (dof) are being used in the solution of some important practical problems. A brief history of the beginning of the finite element method was presented by Gupta and Meek. The rapid progress of the finite element method can be seen by noting that, annually about 3800 papers were being published with a total of about 56,000 papers and 380 books and 400 conference proceedings published as estimated in 1995. With all the progress, today the finite element method is considered one of the well-established and convenient analysis tools by engineers and applied scientists.
Engineering Applications of the Finite Element Method As stated earlier, the finite element method was developed originally for the analysis of aircraft structures. However, the general nature of its theory makes it applicable to a wide variety of boundary value problems in engineering. A boundary value problem is one in which a solution is sought in the domain (or region) of a body subject to the satisfaction of prescribed boundary (edge) conditions on the dependent variables or their derivatives. Table 1-1 gives specific applications of the finite element in the three major categories of boundary value problems, namely (1) equilibrium or steady-state or time-independent problems, (2) eigenvalue problems, and (3) propagation or transient problems. In an equilibrium problem, we need to find the steady-state displacement or stress distribution if it is a solid mechanics problem, temperature or heat flux distribution if it is a heat transfer problem, and pressure or velocity distribution if it is a fluid mechanics problem. In eigenvalue problems also, time will not appear explicitly. They may be considered as extensions of equilibrium problems in which critical values of certain parameters are to be determined in addition to the corresponding steady-state configurations. In these problems, we need to find the natural frequencies or buckling loads and mode shapes if it is a solid mechanics or structures problem, stability of laminar flows if it is a fluid mechanics problem, and resonance characteristics if it is an electrical circuit problem. The propagation or transient problems are time-dependent problems. This type of problem arises, for example, whenever we are interested in finding the response of a body under timevarying force in the area of a solid mechanics and under sudden heating or cooling in the field of heat transfer.
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Commercial Finite Element Program Packages The general applicability of the finite element method makes it a powerful and versatile too l for a wide range of problems. Hence, a number of computer program packages have been developed for the solution of a variety of structural and solid mechanics problems. Some of the programs have been developed in such a general manner that the same program can be used for the solution of problems belonging to different branches of engineering with little or no modification. Many of these packages represent large programs that can be used for solving real complex problems. For example, the NASTRAN (National Aeronautics and Space Administration Structural Analysis) program package contains approximately 150,000 FORTRAN statements and can be used to analyze physical problems of practically any size, including a complete aircraft, an automobile, and a space shuttle.
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The availability of supercomputers has made a strong impact on the finite element technology. In order to realize the full potential of these supercomputers in finite element computation, special parallel numerical algorithms, programming strategies, and programming languages are being developed. The use of personal computers and workstations in engineering analysis and design is becoming increasingly popular as the price of hardware is decreasing dramatically. Many finite element programs, especially suitable for the personal computer and workstation environment, have been developed. Among the main advantages are a user-friendly environment and inexpensive graphics.
Solutions Using Finite Element Software Three Steps of Finite Element Solution The solution of any engineering analysis problem us ing commercial FEA software involves the following three steps: Preprocessing: In this step, the geometry, material properties, loads (actions) and boundary conditions are given as input data. In-built automatic mesh generation modules develop the finite element mesh with minimal input from the analyst on the type of elements and mesh density to be used. The analyst can display the data as well as the finite element mesh generated for visual inspection and verification for correctness. Numerical analysis: The software automatically generates the element characteristics (stiffness) matrices and characteristic (load) vectors, assembles them to generate the system equations, implements the specified boundary conditions and solves the equations to find the nodal values of the field variable (displacements) and computes the element resultants (stresses and strains). Post processing: The solution of the problem, such as nodal displacements and element stresses, can be displayed either numerically in tabular form or graphically (two- or three-dimensional plots of deformed shape or stress variation. The analyst can choose the mode of display for the results. Checking the Results of FEA It is extremely important to check the results given by the FEA software. Usually a simpler version of the actual problem is to be solved using the software so that the results can be compared with known solutions (obtained by other methods such as a simplified analysis technique). In addition, the analyst must ensure that the results agree with engineering intuition and behavior. Also, one needs to verify whether the solution satisfies the specified boundary and symmetry conditions. If necessary, the problem needs to be so lved by changing the boundary conditions, loads or materials to find whether the resulting FEA solutions behave as per engineering intuition and expectations.
Basic Element Shapes In most engineering problems, we need to find the values of a field variable such as displacement, stress, temperature, pressure, and velocity as a function of spatial coordinates
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(x, y, z). In the case of transient or unsteady-state problems, the field variable has to be found as a function of not only the spatial coordinates (x, y, z) but also time (t). The geometry (domain or solution region) of the problem is often irregular. The first step of the finite element analysis involves the discretization of the irregular domain into smaller and regular subdomains, known as finite elements. This is equivalent to replacing the domain having an infinite number of degrees of freedom (dof) by a sy stem having a finite number of dof. A variety of methods can be used to model a domain with finite elements . Different methods of dividing the domain into finite elements involve varying amounts of computational time and often lead to different approximations to the solution of the physical problem. The process of discretization is essentially an exercise of engineering judgment. Efficient methods of finite element idealization require some experience and knowledge of simple guidelines. For large problems involving complex geometries, finite element idealization based on manual procedures requires considerable effort and time on the part of the analyst. Some automatic mesh generation programs have been developed for the efficient idealization of complex domains requiring minimal interface with the analyst. The shapes, sizes, number, and configurations of the elements have to be chosen carefully such that the original body or domain is simulated as closely as possible without increasing the computational effort needed for the solution. Mostly the choice of the ty pe of element is dictated by the geometry of the body and the number of independent coordinates necessary to describe the system. If the geometry, material properties, and the field variable of the problem can be described in terms of a single spatial coordinate, we can use the onedimensional or line elements shown in Figure 1-4(a). The temperature distribution in a rod (or fin), the pressure distribution in a pipe flow, and the deformation of a bar under axial load, for example, can be determined using these elements. Although these elements have a crosssectional area, they are generally shown schematically as a line element (Figure 1-4(b)).In some cases, the cross-sectional area of the element may be non-uniform.
Figure 1-4
For a simple analysis, one-dimensional elements are assumed to have two nodes, one at each end, with the corresponding value of the field variable chosen as the unknown (degree of freedom). However, for the analysis of beams, the values of the field variable (transverse displacement) and its derivative (slope) are chosen as the unknowns (dof) at each node as shown in Figure 1-4(c).
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When the configuration and other details of the problem can be described in terms of two independent spatial coordinates, we can use the two-dimensional elements shown in Figure 1-5. The basic element useful for two-dimensional analysis is the triangular element. Although a quadrilateral element (or its special forms, the rectangle and parallelogram) can be obtained by assembling two or four triangular elements, as shown in Figure 1-6, in some cases the use of quadrilateral (or rectangle or parallelogram) elements proves to be advantageous. For the bending analysis of plates, multiple dof (transverse displacement and its derivatives) are used at each node.
Figure 1-5
Figure 1-6
If the geometry, material properties, and other parameters of the body can be described by three independent spatial coordinates, we can idealize the body by using the threedimensional elements shown in Figure 1-7. The basic three-dimensional element, analogous
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to the triangular element in the case of two-dimensional problems, is the tetrahedron element. In some cases the hexahedron element, which can be obtained by assembling five tetrahedrons as indicated in Figure 1-7, can be used advantageously. Some problems, which are actually three-dimensional, can be described by only one or two independent coordinates. Such problems can be idealized by using an axisymmetric or ring type of elements shown in Figure 1-8. The problems that possess axial symmetry, such as pistons, storage tanks, valves, rocket nozzles, and reentry vehicle heat shields, fall into this category.
Figure 1-7
Figure 1-8
For the discretization of problems involving curved geometries, finite elements with curved sides are useful. Typical elements having curved boundaries are shown in Figure 1-9. The ability to model curved boundaries has been made possible by the addition of mid-side nodes. Finite elements with straight sides are known as linear elements, whereas those with curved sides are called higher-order elements.
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Figure 1-9
Discretization Process Various considerations to be taken in the discretization process are discussed in the following sections.
Type of Elements Often, the type of elements to be used will be evident from the physical problem. For example, if the problem involves the analysis of a truss structure under a given set of load conditions (Figure 1-10(a)), the type of elements to be used for idealization is obviously the “bar or line elements” as shown in Figure 1-10(b). Similarly, in the case of stress analysis of the short beam shown in Figure 1-11(a), the finite element idealization can be done using three-dimensional solid elements as shown in Figure 1-11(b). However, the type of elements to be used for idealization may not be apparent, and in such cases one has to choose the ty pe of elements judicially. As an example, consider the problem of analysis of the thin-walled shel l shown in Figure 1-12(a). In this case, the shell can be idealized by several types of elements as shown in Figure 1-12(b). Here, the number of dof needed, the expected accuracy, the ease with which the necessary equations can be derived, and the degree to which the physical structure can be modeled without approximation will dictate the choice of the element type to be used for idealization. In certain problems, the given body cannot be represented as an assemblage of only one type of elements. In such cases, we may have to use two or more types of elements for idealization. An example of this would be the analysis of an aircraft wing. Since the wing consists of top and bottom covers, stiffening webs, and flanges, three types of elements —namely, triangular plate elements (for covers), rectangular shear panels (for webs), and frame elements (for flanges) —have been used in the idealization shown in Figure 1-13.
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Figure 1-10
Figure 1-11
Figure 1-12
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Figure 1-13
Size of Elements The size of elements influences the convergence of the solution directly, and hence it has to be chosen with care. If the size of the elements is small, the final solution is expected to be more accurate. However, we have to remember that the use of smaller-sized elements will also mean more computation time. Sometimes, we may have to use elements of different sizes in the same body. For example, in the case of stress analysis of the box beam shown in Figure 1-14(a), the size of all the elements can be approximately the same, as shown in Figure 1-14(b). However, in the case of stress analysis of a plate with a hole shown in Figure 1-15(a), elements of different sizes have to be used, as shown in Figure 1-15(b). The size of elements has to be very small near the hole (where stress concentration is expected) compared to distant places. In general, whenever steep gradients of the field variable are expected, we have to use a finer mesh in those regions. Another characteristic related to the size of elements that affects the finite element solution is the aspect ratio of the elements. The aspect ratio describes the shape of the element in the assemblage of elements. For twodimensional elements, the aspect ratio is taken as the ratio of the largest dimension of the element to the smallest dimension. Elements with an aspect ratio of nearly unity generally yield best results.
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Figure 1-14
Figure 1-15
Location of Nodes If the body has no abrupt changes in geometry, material properties, and external conditions (e.g., load and temperature), the body can be divided into equal subdivisions and hence the spacing of the nodes can be uniform. On the other hand, if there a re any discontinuities in the problem, nodes have to be introduced at these discontinuities, as shown in Figure 1 -16.
Figure 1-16
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Number of Elements The number of elements to be chosen for idealization is related to the accuracy desired, siz e of elements, and the number of dof involved. Although an increase in the number of elements generally means more accurate results, for any given problem, there will be a certain number of elements beyond which the accuracy cannot be significantly improved. This behavior is shown graphically in Figure 1-17. Moreover, since the use of a large number of elements involves a large number of dof, we may not be able to store the resulting matrices in the available computer memory.
Figure 1-17
Simplifications Afforded by the Physical Configuration of the Body If the configuration of the body as well as the external conditions are symmetric, we may consider only half of the body for finite element idealization. The symmetry conditions, however, have to be incorporated in the solution procedure. This is illustrated in Figure 1-18, where only half of the plate with a hole, having symmetry in both geometry and loading, is considered for analysis. 1 Since there cannot be a horizontal displacement along the line of symmetry AA, the condition that u=0 has to be incorporated while finding the solution.
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Figure 1-18
Finite Representation of Infinite Bodies In most of the problems, like in the analysis of beams, plates, and shells, the boundaries of the body or continuum are clearly defined. Hence, the entire body can be considered for element idealization. However, in some cases, as in the analysis of dams, foundations, and semi-infinite bodies, the boundaries are not clearly defined. In the case of dams (Figure 1-19), since the geometry is uniform and the loading does not change in the length direction, a unit slice of the dam can be considered for idealization and analyzed as a plane strain problem. However, in the case of the foundation problem shown in Figure 1-20(a), we cannot idealize the complete semi-infinite soil by finite elements. Fortunately, it is not really necessary to idealize the infinite body. Since the effect of loading decreases gradually with increasing distance from the point of loading, we can consider only that much of the continuum in which the loading is expected to have a significant effect as shown in Figure 1-20(b). Once the significant extent of the infinite body is identified as shown in Figure 1-20(b), the boundary conditions for this finite body have to be incorporated in the solution. For example, if the horizontal movement only has to be restrained for sides AB and CD (i.e., u=0), these sides are supposed to be on rollers as shown in Figure 1-20(b). In this case, the bottom boundary can be either completely fixed (u=v=0) or constrained only against vertical movement (v=0). The fixed conditions (u=v=0alongBC) are often used if the lower boundary is taken at the known location of a bedrock surface.
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Figure 1-19
Figure 1-20
NODE NUMBERING SCHEME The finite element analysis of practical problems often leads to matrix equations in which the matrices involved will be banded. The advances in the finite element analysis of large practical systems have been made possible largely due to the banded nature of the matrices. Furthermore, since most of the matrices involved (e.g., stiffness matrices) are symmetric, the demands on the computer storage can be substantially reduced by storing only the elements involved in half bandwidth instead of storing the entire matrix. The bandwidth of the overall or global characteristic matrix depends on the node numbering scheme and the number of dof considered per node. If we can minimize the bandwidth, the storage requirements as well as solution time can also be minimized. Since the number of dof per node is generally fixed for any given type of problem, the bandwidth can be minimized by using a proper node numbering scheme. As an example, consider a three-bay frame with rigid joints, 20 stories high, shown in Figure 1-21. Assuming that there are 3 dof per node, there are 252 unknowns in the final equations (including the dof corresponding to the fixed nodes), and if the entire stiffness matrix is stored in the computer, it will require 252 2 =63504 locations. The bandwidth (strictly speaking, half-bandwidth) of the overall stiffness matrix can be shown to be 15, and thus the storage required for the upper half-band is only 15×252=3780 locations.
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Figure 1-21
Before we attempt to minimize the bandwidth, we discuss the method of calculating the bandwidth. For this, we consider again the rigid jointed frame shown in Figure 1.21. By applying constraints to all the nodal dof except number 1 at node 1 (joint A), it is clear that an imposed unit displacement in the direction of 1 will require constraining forces at the nodes directly connected to node A —that is, B and C. These constraining forces are nothing but the cross-stiffnesses appearing in the stiffness matrix, and these forces are confined to the nodes B and C. Thus, the nonzero terms in the first row of the global stiffness matrix (Figure 1.22) will be confined to the first 15 positions. This defines the bandwidth (B)as the maximum difference between the numbered dof at the ends of any membe r + 1. This definition can be generalized so as to be applicable for any type of finite element as Bandwidth (B)=(D+1)*f. Where D is the maximum largest difference in the node numbers occurring for all elements of the assemblage, and f is the number of dof at each node.
Figure 1-21
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The previous equation indicates that D has to be minimized in order to minimize the bandwidth. Thus, a shorter bandwidth can be obtained simply by numbering the nodes across the shortest dimension of the body. This is clear from Figure 1-23 also, where the numbering of nodes along the shorter dimension produces a bandwidth of B=15 (D=4), whereas the numbering along the longer dimension produces a bandwidth of B=66 (D=21). As observed previously, the bandwidth of the overall system matrix depends on the manner in which the nodes are numbered. For simple systems or regions, it is easy to label the nodes so as to minimize the bandwidth. But for large systems, the procedure becomes nearly impossible. Hence, automatic mesh generation algorithms, capable of discretizing any geometry into an efficient finite element mesh without user intervention, have been developed. Most commercial finite element software has built-in automatic mesh generation codes. An automatic mesh generation program generates the locations of the node points and elements, labels the nodes and elements, and provides the element –node connectivity relationships.
Figure 1-23
Automatic Mesh Generation Mesh generation is the process of dividing a physical domain into smaller subdomains (called elements) to facilitate an approximate solution of the governing ordinary or partial differential equation. For this, one-dimensional domains (straight or curved lines) are subdivided into smaller line segments, two-dimensional domains (planes or surfaces) are subdivided into triangle or quadrilateral shapes, and three-dimensional domains (volumes) are subdivided into tetrahedron and hexahedron shapes. If the physical domain is simple and the number of elements used is small, mesh generation can be done manually. However, most practical problems, such as those encountered in aerospace, automobile, and construction industries have complex geometries that require the use of thousands and sometimes millions of elements. In such cases, the manual process of mesh generation is impossible and we have to use automatic mesh generation schemes based on the use of a CAD or solid modeling package.
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Automatic mesh generation involves the subdivision of a given domain, which maybe in the form of a curve, surface, or solid (described by a CAD or solid modeling package) into a set of nodes (or vertices) and elements (subdomains) to represent the domain as closely as possible subject to the specified element shape and size restrictions. Many automatic mesh generation schemes use a “bottom-up” approach in that nodes (or vertices or corners of the domain) are meshed first, followed by curves (boundaries), then surfaces, and finally solids. Thus, for a given geometric domain of the problem, nodes are first placed at the corner points of the domain, and then nodes are distributed along the geometric curves that define the boundaries. Next, the boundary nodes are used to deve lop nodes in the surface(s), and finally the nodes on the various surfaces are used to develop nodes within the given volume (or domain). The nodes or mesh points are used to define line elements if the domain is onedimensional, triangular, or quadrilateral elements if the domain is two-dimensional, and tetrahedral or hexahedral elements if the domain is three-dimensional. The automatic mesh generation schemes are usually tied to solid modeling and computeraided design schemes. When the user supplies information on the surfaces and volumes of the material domains that make up the object or system, an automatic mesh generator generates the nodes and elements in the object. The user can also specify minimum permissible element sizes for different regions of the object. Many mesh generation schemes first create all the nodes and then produce a mesh of triangles by connecting the nodes to form triangles (in a plane region). In a particular scheme, known as Delaunay triangulation, the triangular elements are generated by maximizing the sum of the smallest angles of the triangles; thus the procedure avoids generation of thin elements. The most common methods used in the development of automatic mesh generators are the tesselation and octree methods. In the tesselation method, the user gives a collection of node points and also an arbitrary starting node. The method then creates the first simplex element using the neighboring nodes. Then a subsequent or neighboring element is generated by selectingthe node point that gives the least distorted element shape. The procedure is continued until all the elements are generated. The step-by-step procedure involved in this method is illustrated in Figure 1-24 for a two-dimensional example. Alternately, the user can define the boundary of the object by a series of nodes. Then the tesselation method connects selected boundary nodes to generate simplex elements. The stepwise procedure used 3 in this approach is shown in Figure 1-25.
Figure 1-24
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Figure 1-25
The octree methods belong to a class of mesh generation schemes known as tree structure methods, which are extensively used in solid modeling and computer graphics display methods. In the octree method, the object is first considered enclosed in a three -dimensional cube. If the object does not completely (uniformly) cover the cube, the cube is subdivided into eight equal parts. In the two-dimensional analog of the octree method, known as the quadtree method, the object is first considered enclosed in a square region. If the object does not completely cover the square, the square is subdivided into four equal quadrants. If any one of the resulting quadrants is full (completely occupied by the object) or empty (not occupied by the object), then it is not subdivided further. On the other hand, if any one of the resulting quadrants is partially full (partially occupied by the object), it is subdivided into four quadrants. This procedure of subdividing partially full quadrants is continued until all the resulting regions are either full or empty, or until some predetermined level of resolution is achieved. At the final stage, the partially full quadrants are assumed to be either full or empty arbitrarily based on a pre-specified criterion.
Units Before starting to define any model, you need to decide which system of units you will use. Abaqus has no built-in system of units. Do not include unit names or labels when entering data in Abaqus. All input data must be specified in consistent units. Some common systems of consistent units are shown in table 1-2. Table 1-2
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Chapter 2 Geometric Modeling In this chapter geometric modeling in the part module of the ABAQUS is investigated.
Example 1 Two and three dimensional wire parts
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Example 2 Two dimensional shell parts
Part (1)
Part (2)
Part (3)
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Part (4)
Part (5)
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Part (7)
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Example 3 Three dimensional shell parts
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Example 4 Three dimensional solid parts
Part (1)
Part (2)
Part (3)
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Part (4)
Part (5)
Part (6)
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Part (7)
Part (8)
Part (9)
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Part (10)
Part (11)
Part (12)
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Part (13)
Part (14)
Part (15)
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Part (16)
Part (17)
Part (18)
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Chapter 3 Truss Structures Example 1 The structure is a simple, pin-jointed truss that is constrained at the left end and mounted on rollers at the right end. The members can rotate freely at the joints. The frame is prevented from moving out of plane. A simulation is required to de termine the structure’s static deflection and the peak stress in its members when a 10 kN load is applied as shown in figure. All members are circular steel rods 5 mm in diameter. Elastic properties are E=200 GPa and v=0.29.
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Example 2 E =10e4 ksi, A=2 in 2, P4=P8= 100 kip
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Example 3 E=2×10e 7 N/cm2, A=2cm2 for all members All dimensions in centimeters; All base nodes fixed
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Example 4 Finite element to be used: T3D2 = 3D two-node truss element Dimensions in figure are in mm; Area of cross section of each bar: 3225.8 mm 2 Material: Aluminum; E= 69 GPa Load applied: Vertical load of 10,000 N at node 1
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Example 5 In this example the 52 bar space truss (dome structure) with configuration shown in the following figures is considered. At each free node (1 –13) it is attached a non-structural mass of 50 kg. The material is steel with Young’s modulus equal to 210 GPa , Poisson ration equal to 0.29 and specific mass of 7800 kg/m 3. The 52 bars are divided into eight groups, as shown in the table.
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Other Examples
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Chapter 4 Two Dimensional Elasticity Example 1 Hole in Plate A Rectangular plate with central circular hole is subjected to a uniformly distributed axial force as shown. If the plate was made of carbon steel with 210 GPa Young modulus and 0.29 Poisson ratio determine the maximum deflection and Von-Mises stress due to the specified loading. The plate thickness is 0.002 and the resultant of the axial loading is 2 kN. Use any symmetry in the model if it is appropriate. Also do a grid study analysis and determine the optimal mesh size.
50 45 40
) a 35 p M ( 30 s s e 25 r t S m20 u m15 i x a 10 M
S (Mpa) Q4 S (Mpa) T3
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Example 2 L Shaped Bracket An L shaped plate of thickness 0.05 inch is subjected to a distributed force as shown. If the plate was made of aluminum with 10.4 Mpsi Young modulus and 0.333 Poisson ratio determine the maximum deflection and Von-Mises stress due to the specified loading.
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Example 3 Tensile Bracket A bracket plate with an internal hole is loaded axially. The plate was made of carbon steel with 210 GPa Young modulus and 0.29 Poisson ratio. Determine the maximum deflection and Von-Mises stress due to an axial tensile loading of 10 kN distributed along the right and left edges.
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Example 4 Connecting Rod Consider the following connecting rod (All dimensions are in mm). Assume the left hole constrained completely in all directions and 1 KN vertical force is distributed over the internal surface of the right hole. If it was 10mm in thickness, determine the maximum Von-mises stress and maximum displacement for the plane-stress case. The rod is made of ST37 structural steel with 200GPa Young modulus and 0.3 Poisson ratio.
The solution is as follows. All deformations are in mm and all stresses are in MPa.
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Example 5 Curved Slider Consider the following slider (all dimensions are in inch). The left hole is constrained in all directions and the right hole constrained only in y direction. Also assume that 250 lb force is distributed in the semicircular arc at the upper end of the slit with an angle of 30 degree with respect to the horizon. If the thickness of the plate was 1/8 inch and it was made of structural steel with 30Gpsi Young modulus and 0.3 Poisson ratio determine the maximum Von-mises stress and also the maximum deflection in the object.
The solution is as follows. All deformations are in inch and all stresses are in psi.
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Example 6 Concentrated Force Consider a 2D square object supported on a flat frictionless surface as shown in the following figure. A concentrated force is applied at the midpoint of its upper edge. Use linear and quadratic quadrilateral elements to evaluate the maximum Von-Misses stress under the load. Examine different mesh size and draw convergence curve. The plate is made from structural steel with E=200GPa and v=0.3. Its thickness is 1mm and assume plane stress.
180 160 Element Q4 140
Element Q8
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Example 7 Sharp Corner Consider a 2D object with sharp corners which is clamped at left and is under a uniform tensile stress at right as shown in the following figure. Use linear and quadratic quadrilateral elements to evaluate the maximum Von-Misses stress in the object. Examine different mesh size and draw convergence curve. The plate is made from structural steel with E=200GPa and v=0.3. Its thickness is 1mm and assume plane stress.
Q4 Elements
Q8 Elements
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2.8 2.6
Element Q4 (Mpa)
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) a p M2.2 ( s s 2 e r t S m1.8 u m i x 1.6 a M
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Example 8 Rounded Corner Consider a 2D object with round corners which is clamped at left and is under a uniform tensile stress at right as shown in the following figure. Use linear and quadratic quadrilateral elements to evaluate the maximum Von-Misses stress in the object. Examine different mesh size and draw convergence curve. The plate is made from structural steel with E=200GPa and v=0.3. Its thickness is 1mm and assume plane stress.
Q4 Elements
Q8 Elements
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1.3
1.25 ) a p M1.2 ( s s e r t1.15 S m u m i x 1.1 a M
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Example 9 Solid Slab under Variable Load Distribution A long rectangular solid slab is clamped completely along two opposite sides as shown in the figure. The slab is subjected to a transvers triangular distributed force with maximum of 1 00 tonne/m for unit depth. The slab was made of concrete with 25 GPa Young modulus, 0.2 Poisson ratio and 2320 kg/m 3 mass density. Determine the maximum deflection and maximum normal stress in the slab due to the specified loading.
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Example 10 Solid Slab under Variable Load Distribution
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Example 11 Pipe Made of Two Different Materials
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Chapter 5 Three Dimensional Elasticity Example 1 Cantilever Beam
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Example 2 Cantilever Beam under Variable Loading
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Example 3 L Shaped Bracket 1 Consider the following 3D bracket (all dimensions are in mm). Assume that the left end of the object is clamped completely and a 1KN shearing force is distributed over the internal surface of the hole and is pointing to the right. The bracket is made of structural steel with 200GPa Young modulus and 0.3 Poisson ratio. In this circumstance, determine the maximum Vonmises stress and maximum displacement produced in the object.
The solution is as follows. All deformations are in mm and all stresses are in MPa.
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Example 4 L Shaped Bracket 2 Consider the following 3D bracket (all dimensions are in mm). Assume that the internal surfaces of the two holes in the right leg are clamped completely and a 3KN shearing force i s distributed over the internal surface of the upper hole and is p ointing to the right. The bracket is made of structural steel with 200GPa Young modulus and 0.3 Poisson ratio. In this circumstance, determine the maximum Von-Mises stress and maximum displacement produced in the object.
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Example 5 Pin Loading
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Chapter 6 Heat Transfer Example 1 Conductive Heat Transfer in a Square Three sides of a square plate are maintained at constant temperature of 0 C and the fourth one is kept at 100 C. The plate is made of carbon steel with average thermal conductivity of 52 W/mK. Under this conditions, determine the temperature distribution in the plate. °
°
K=(53.2+50.7)/2=52 W/mK
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100.00
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Example 2 Three Dimensional Needle Fin Consider a needle fin with circular cross section made of carbon steel which is under natural convection in the quiet air. The base of the fin is maintained at 100 C while the ambient temperature is 25 C. Under this condition, determine the temperature distribution along the fin. °
°
The thermal conductivity of the fin can be approximated for 50 C as fillows. °
k=(53.2+50.7)/2=52 W/mK
The coefficient of convective heat transfer can be approximated for horizontal cylinders as the follows: h (1.79 0.179 t
0.25
0.25
h (2.16 0.32 t
t ) d o
)t
1/ 3
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W/m2K
for laminar conditions
W/m2K
for turbulent conditions
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25 , t 100 25 75 and d o
0.01 therefore one can obtain
12.9
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Chapter 7 Thermal Stress Example 1 Thermal Stress in a Rod A steel link, with no internal stresses, is pinned between two solid structures at a reference temperature of 0°C (273 K). One of the solid structures is heated to a temperature of 75°C (348 K). As heat is transferred from the solid structure into the link, the link will attempt to expand. However, since it is pinned, this cannot occur, and as such, stress is created in the link. A steady-state solution of the resulting stress will be found to simplify the analysis. Loads will not be applied to the link, only a temperature change of 75°C. The link is steel with a modulus of elasticity of 200 GPa, a thermal conductivity of 60.5 W/mK, and a thermal expansion coefficient of 12e -6/K.
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Example 2 Thermal Stress in a Square Consider a square 2D elastic object which is under thermal stress condition. Three sides of the square are maintained in constant temperature of 0 C and the fourth one is kept in 100 C as shown in the following figure. Under this condition, determine the temperature distribution in the body and also obtain the maximum displacement and Von-Mises stress induced in it for both cases of plane stress and plane strain. The square is made of carbon steel with the following material properties. °
Coefficient of thermal conductivity: Coefficient of thermal expansion: Young modulus: Poisson ratio:
52 10.8e-6 200e9 0.3.
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W/mK 1/ C Pa °
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Plane stress case
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Plane strain case
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Example 3 Thermal Stress in a Pipe Consider a circular cross section long pipe with internal diameter of 60mm and external diameter of 100mm. The internal surface of the pipe is maintained in 100 C and its external surface is kept in 0 C. Under this condition, determine the temperature distribution, maximum displacement and Von-Mises stress induced in the pipe due to thermal stress. The pipe is made of carbon steel with the following material properties. Coefficient of thermal conductivity: 52 W/mK Coefficient of thermal expansion: 10.8e-6 1/ C Young modulus: 200e9 Pa Poisson ratio: 0.3. °
°
°
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Chapter 8 Beam Structures Example 1 Beam Bridge Young’smodulus: E=70e3 Poisson’s ratio: v=0.0 it is not required for beams
THE FOLLOWING TABLE IS PRINTED FOR NODES BELONGING TO NODES
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Example 2 Beam Bridge The two-dimensional bridge structure is simply supported at its lower corners. The structure is composed of steel T-sections (E=210 GPa, ν=0.25) oriented as shown below. The detail of the cross section is also shown. A uniform distributed load of 1000 N /m is applied to the lower horizontal members in the vertical downward direction. Determine the stresses and the vertical displacements.
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Example 3 Table The following table frame structure is supported by a frictionless surface under its legs. The structure is composed of steel box sections (E=210 GPa, ν=0.25). A uniform distributed load of 1000 N/m is applied to the upper horizontal members in the v ertical downward direction. Determine the stresses and the displacements of the structure.
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Example 4 Crane The following structure is a part of a crane which is clamped at four left joints. The main beams of the structure are composed of steel circular cross section pipes with 80mm outer diameter and 4mm in thickness. The braces are also made of steel pipes of 60mm outer diameter and 3mm in thickness. This structure, In addition to the weights of its own members, supports a concentrated force of 6KN at the mid span of the last member on the right. Under this circumstance, determine the stresses and displacements of the structure. For steel, consider E=210 GPa, ν=0.25 and mass density as 7800 kg/m3. The gravitational acceleration is 9.81m/s2.
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Example 5 Bridge To deal with a more realistic example, the following discussion considers the arch structure shown in the figure, a simplified model of a pedestrian bridge with the arches in a baskethandle configuration. The span L=1200 in (30.5 m); the width W=96 in ( 2.44 m). Cross sections are selected either from a list of 32 AISC HSS round sections or a list of 75 AISC HSS rectangular tube sections. The lists are compiled by selecting the lightest section for each shape variety. The model includes one geometric decision variable, the span-to- depth ratio of the arch (γ), which ranges from 4 to 12, plus the five section variables indicated in Fig.5as follows: Rib: The main arch rib, a round HSS section. Brace: Members that connect the two arches near the crown, a round HSS section. Hanger: The suspenders which transfer load from the deck to the arch, a round HSS section. Tie beam: The longitudinal beams at the deck level, a rectan-gular HSS section. Transverse beam: The beams which span between the tie, a rectangular HSS section. Concerning structural modeling, the structure uses a pin sup-port for each arch at one end of the bridge and roller supports at the other. The hanger members are pin-ended, and the connections between the arches and the tie beams are also pinned. Concerning materials, the rectangular tube sections use a yield stress of 46 ksi (317 MPa), and the round sections use a yield stress of 42 ksi (290 MPa). Dead load includes the self weight of the model plus a superimposed dead load on the deck area of 0:0 8 kip=ft 2 (3.83 kPa). The live load is 0:85 k=ft 2 (4.07 kPa) distributed on the deck area. Superimposed dead and live loads are applied to the nodes of each transverse beam, at the beam ends, and a midspan node according to tributary area. The vertical deflection of the nodes of the tie beam are limited toL=1;000 for live load only. The analysis includes four load combinations: two to check stiffness and two to check strength and stability. The combinations for stiff-ness include one with full live load and one with live load on half the span; these combinations use linear analysis. The combinations for strength and stability include one with dead plus live and an-other with dead plus l ive load on half the span; these combinations are factored according to the AISC LRFD code (AISC 2001) and use large displacement analysis. Note that these load combinations are unrealistically simple, in particular because they do not account for lateral loads. Concerning stability criteria, the algorithm checks each member to account for member-level compression stability according to the AISC LRFD requirements discussed previously by using an effective length factorK¼1:0 and an unsupported 78
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length equal to the member length. System-level stability was considered as described in the discussion of stability constraints
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Chapter 9 Shell and Plate Structures Example 1 Rectangular Plate Plate is fixed at one edge and supported by rollers at the opposite edge. Other two e dges are free (no support). A concentrated transverse force of 100 N applied at center. Material: E=10e3 N/mm2, Poisson ratio = 0.3
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Example 2 L Shaped Beam Structure A frame structure which is composed of two I beams is shown in the following. The dimensions of the cross section are also shown. Assume the structure is completely fixed at point A and a vertical force of 1KN is applied at the end at B. It is also needed to consider the weight of the own structure. The structure is made of structural steel with 200GPa Young modulus, 0.3 Poisson ratio and 7800kg/m3 mass density. In these conditions determine stress distribution and displacements of the structure.
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Example 3 Pressure Vessel A pressure vessel is composed of a cylindrical body with hemispherical caps. The diameter of the vessel is 1m and its total length is 2m. Two pipes of 0.2m diameter are attached to the central part of the vessel with relative angle of 90 degrees. Each pipe is 0.3m in length and the distance between centerlines is 0.6m. The connection point of the pipes and the vessel is filleted with radius of 0.05m. The thickness of the ve ssel and pipes is 0.005m and are made of structural steel (E=200GPa, v=0.3). Assume a constant pressure of 2atm is exerted on the internal surfaces and the free ends of the pipes are clamped completely. In these conditions determine the stress and displacement distribution in the vessel.
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Example 4 Pipe E=200GPa Niu=0.3 Thichness=0.01 m Internal Pressure=1e6 Pa BC: onlt the translational DOFs of the edges of the bolt holes are restrained
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Chapter 10 Composite Structures Composite is a macroscopic mixture of at least two materials. One of the materials is the matrix in which the other materials called reinforcements are embedded. The following figure shows a schematic of an Airbus 380 airplane (the largest airplane in the world as of 2008). This airplane has more than 50% of its structure made of composite materials. These components include the flaps, ailerons, rudder, radome etc. Most of these components are flat in shape and they are usually made using hand-lay-up (HLU) and autoclave molding techniques. The next figure shows a schematic of the hand-lay-up fabrication technique and a representative lay-up sequence. Autoclave molding is a wellestablished method for composites used in the aero-space industry with certified resins and fibers. A photograph of an auto-clave is also shown.
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The following figure shows a pressure vessel made of composite materials using the combination of hand-lay-up and filament winding processes. Composite pressure vessels are light weight and can contain pressures higher than those contained by metallic vessels. These components are made using the filament winding process.
The following figure shows a photograph of a filament winding machine.
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Figure 1.3(a) shows a component made using pultrusion. Pultrusion is used to make many structures for civil engineering applications. Figure 1.3(b) shows the schematic of the pultrusion process, and Figure 1.3(c) shows a photograph of a lab scale pultrusion machine.
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Figure 1.4(a) shows a composite component made using the liquid composite molding (LCM) method (5 piece). LCM has been used to make automobile composite components. Figure 1.4(b) shows a schematic of the liquid composite molding process.
Depending on the purpose of the analysis, different modeling techniques for composites can be used: Microscopic modeling Matrix and reinforcements are separately modeled as deformable continua. Each element is composed of a single homogeneous material. Layered modeling Each element is composed of several layers of different materials. Smeared modeling The composite is modeled as an equivalent homogeneous material with stacked or single layer element configuration
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Example 1 Glass/Epoxy E1=38.6 GPa E2=8.27 GPa G12=4.14 GPa Viu12=0.26 Layer thickness =0.002 m Layup: [90, 45, 0, 45, 90]
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Chapter 11 Free Vibration
Example 1
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Chapter 12 Linear Buckling
Example 1
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Chapter 13 Contact Stress Example 1
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Chapter 14 Plastic Deformation
Example 1
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Chapter 15 Flow in Porous Media
Example 1
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Chapter 16 Flow of Viscose Fluids
Example 1
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Appendix 1 Material Properties
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Appendix 2 Stress Concentration Factors
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Appendix 3 Meshing Techniques Top-down meshing Top-down meshing relies on the geometry of a part to define the outer bounds of the mesh. The top-down mesh matches the geometry; you may need to simplify and/or partition complex geometry so that Abaqus/CAE recognizes basic shapes that it can use to generate a high-quality mesh. In some cases top-down methods may not allow you to mesh portions of a complex part with the desired type of elements. The top-down techniques —structured, swept, and free meshing —and their geometry requirements are well-defined, and loads and boundary conditions applied to a part are associated automatically with the resulting mesh. Structured meshing Structured meshing is the top-down technique that gives you the most control over your mesh because it applies preestablished mesh patterns to particular model topologies. Most unpartitioned solid models are too complex to be meshed using preestablished mesh patterns. However, you can often partition complex models into simple regions with topologies for which structured meshing patterns exist. Figure 17 –3 shows an example of a structured mesh.
Swept meshing Abaqus/CAE creates swept meshes by internally generating the mesh on an edge or face and then sweeping that mesh along a sweep path. The result can be either a two-dimensional mesh created from an edge or a three-dimensional mesh created from a face. Like structured meshing, swept meshing is a top-down technique limited to models with specific topologies and geometries. Figure 17 –4 shows an example of a swept mesh.
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Free meshing The free meshing technique is the most flexible top-down meshing technique. It uses no preestablished mesh patterns and can be applied to almost any model shape. However, free meshing provides you with the least control over the mesh since there is no way to predict the mesh pattern. Figure 17 –5 shows an example of a free mesh.
Bottom-up meshing Bottom-up meshing uses the part geometry as a guideline for the outer bounds of the mesh, but the mesh is not required to conform to the geometry. Removing this restriction gives you greater control over the mesh and allows you to create a hexahedral or hexahedraldominated mesh on geometry that is too complex for the structured or swept meshing techniques. Bottom-up meshing can be applied to any solid model shape. It provides you with the most control over the mesh, since you select the method and the parameters that drive the mesh. However, you must also decide whether the resulting mesh is a suitable approximation of the geometry. If it is not, you can delete the mesh and try a different bottom-up meshing method or partition the region and mesh the resulting smaller regions with either bottom-up or top-down meshing techniques. To mesh a single bottom-up region, you may have to apply several successive bottom-up meshes. For example, you may use an extruded bottom-up mesh to generate part of a region,
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then use the element faces of the extruded mesh as a starting point to generate a swept mesh for features that the extruded mesh did not include. Loads and boundary conditions are applied to geometry. Unlike a top-down mesh, a bottomup mesh may not be fully associated with geometry. Therefore, you should check that the mesh is correctly associated with the geometry in areas where loads or boundary conditions are applied. Proper mesh-geometry association will ensure that the loads and boundary conditions are correctly transferred to the mesh during the analysis. (For more information, see “Mesh-geometry association,” Section 17.11.4.) Because of the extra effort requi red by the user to create a satisfactory mesh compared to the automated top-down meshing processes, bottom-up meshing is recommended for use only when top-down meshing cannot generate a suitable mesh. Figure 17 –6 shows an example of a bottom-up meshed part. Although this part is relatively simple, it requires two regions and four bottom-up meshes to completely mesh the part. Abaqus/CAE displays bottom-up meshed regions using a mixture of the region geometry color (light tan) and the mesh color (light blue) to emphasize that the geometry and mesh may not be associated. Displaying both the geometry and the mesh allows you to view and edit the mesh-geometry associativity.
What is structured meshing? The structured meshing technique generates structured meshes using simple predefined mesh topologies. Abaqus/CAE transforms the mesh of a regularly shaped region, such as a square or a cube, onto the geometry of the region you want to mesh. You can apply the structured meshing technique to simple two-dimensional regions (planar or curved) or to simple three-dimensional regions that have been assigned the Hex or Hex-dominated element shape option. For example, Figure 17 –42 illustrates how simple mesh patterns for triangles, squares, and pentagons are applied to more complex shapes.
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Three-dimensional structured meshing Figure 17 –49 illustrates examples of simple three-dimensional regions that can be meshed using the structured meshing technique.
Meshing more complex regions with this technique may require manual partitioning. If you do not partition a complex region, your only meshing option may be the free meshing technique with tetrahedral elements. Meshes constructed using the structured meshing technique consist of hexahedral elements, which are preferred over tetrahedral elements. You can eliminate holes (whether they pass all the way through the part instance or just part way through) by partitioning their circumferences into halves, quarters, etc. For example, the four partitions in Figure 17 –51 convert the part instance from one region with a hole to four regions without holes.
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M.J. Kazemzadeh-Parsi
You should limit arcs to 90° or less to avoid concavities along sides and at edges. For example, the part instance in Figure 17 –52 has been partitioned so that the single region with 180° arcs becomes two regions with 90° arcs.
All the faces of the region must have geometries that could be meshed using the twodimensional structured meshing technique. For example, without partitioning, the semicircles at either end of the part in Figure 17 –53 have only two sides each. (A face must have at least three sides to be meshed using the structured meshing technique.) If you partition the part into two halves, each semicircle is divided into two faces with three sides each.
Exactly three edges of the region must meet at each vertex. For example, the vertex at the top of an unpartitioned pyramid in Figure 17 –54 is connected to four edges. However, if you partition the pyramid into two tetrahedral regions, the vertex is connected to only three edges for each individual region.
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ABAQUS Lecture Notes
M.J. Kazemzadeh-Parsi
What is swept meshing? Abaqus/CAE uses swept meshing to mesh complex solid and surface regions. The swept meshing technique involves two phases:
Abaqus/CAE creates a mesh on one side of the region, known as the source side. Abaqus/CAE copies the nodes of that mesh, one element layer at a time, until the final side, known as the target side, is reached. Abaqus/CAE copies the nodes along an edge, and this edge is called the sweep path. The sweep path can be any type of edge —a straight edge, a circular edge, or a spline. If the sweep path is a straight edge or a spline, the resulting mesh is called an extruded swept mesh. If the sweep path is a circular edge, the resulting mesh is called a revolved swept mesh. For example, Figure 17 –65 shows an extruded swept mesh. To mesh this model, Abaqus/CAE first creates a two-dimensional mesh on the source side of the model. Next, each of the nodes in the two-dimensional mesh is copied along a straight edge to every layer until the target side is reached.
To determine if a region is swept meshable, Abaqus/CAE tests if the region can be replicated by sweeping a source side along a sweep path to a target side. In general, Abaqus/CAE selects the most complex side (for example, the side that has an isolated edge or vertex) to be the source side. In some cases you can use the mesh controls to select the sweep path. If some regions of a model are too complex to be swept meshed, Abaqus/CAE asks if you want to remove these regions from your selection before it generates a swept mesh on the remaining regions. You can use the free meshing technique to mesh the complex regions, or you can partition the regions into simplified geometry that can be structured or swept meshed. When you assign mesh controls to a region, Abaqus/CAE indicates the direction of the sweep path and allows you to control the direction. If the region can be swept in more than one direction, Abaqus/CAE may generate a very different two -dimensional mesh on the faces that
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